Screw rotor machine for elastic working fluid



Nov. 8, 1966 L. B. SCHIBBYE 3,

SCREW ROTOR MACHINE FOR ELASTIC WORKING FLUID 4 Sheets-Sheet 1 FiledSept. 8, 1964 mvemo Nov. 8, 1966 L. B. SCHIBBYE 3,233,996

SCREW ROTOR MACHINE FOR ELASTIC WORKING FLUID Filed Sept. 8, 1964 4Sheets-Sheet 2 Fig.4

Fig.5

z INVEZIFOR ATTORNEY Nov. 8, 1966 B. SCHIBBYE 3,283,996

SCREW ROTOR MACHINE FOR ELASTIC WORKING FLUID- Filed Sept. 8, 1964 4Sheets-Sheet 3' INVENTOR Nov. 8, 1966 3,283,996

SCREW ROTOR MACHINE FOR ELASTIC WORKING FLUID Filed Sept. 8, 1964 L. B.SCHIBBYE 4 Sheets$heet 4 INVENTOR United States Iatent C 3,283,996 SCREWROTOR MACHINE FOR ELASTIC WORKING FLUID Lauritz Benedictus Schibbye,Saltsjo-Duvnas, Sweden, as-

signor to Svenska Rotor Maskiner Aktiebolag, Nacka, Sweden, acorporation of Sweden Filed Sept. 8, 1964, Ser. No. 394,985

Claims priority, application Sweden, Sept. 12, 1963,

9,983/ 63 9 Claims. (Cl. 230-443) A screw rotor machine for elasticworking fluid is a machine which primarily comprises at least twocoplanar, cooperating rotors provided with helical lands with a wrapangle of less than 360 and intervening grooves, and a housing enclosingthe rotors. The rotors are of male and female type which means that thelands of the male rotors have at least their major portions outside thepitch circle of the rotor and have substantially conyex flanks, and thatthe lands of the female rotor have at least their major portions insidethe pitch circle of the rotor and have substantially concave flanks. Thehousing enclosing the rotors is further provided with a working spacesubstantially composed of two coplanar intermeshing bores, eachsealingly enclosing a rotor, and provided with spaced low pressure andhigh pressure ports, of which at least their major portions are located,respectively, on opposite sides of the plane of the axes of the bores.

The rotors cooperate in such a way that a land of one rotor meshes witha groove of the cooperating rotor on the same side of the plane of theaxes of the bores as the high pressure port is located whereby achevronshaped chamber is formed between the rotors and the walls of theworking space comprising a portion of a male rotor groove and a portionof a female rotor groove communicating therewith. Each such chamber hasits base ends located in a stationary transverse plane common for allchevron-shaped chambers and located at the high pressure port and itstip at the intermesh between the rotor lands of the different rotors.When the rotors revolve the tip of the chevron-shaped chamber is thenmoved axially in relation to the stationary plane so that the volume ofthe chamber is varied.

By variation of the shape of the ports, especially that of the highpressure port, the ratio between the volumes of the chevron-shapedchambers at the moment when the communication with the ports is brokenand opened, respectively, the built-in volume ratio can be varied. Inthis way the ratio between the pressures in the chevronshaped chambersat said moment, the built-in pressure ratio, can also be varied. Thebuilt-in pressure ratio is, however, in addition to the built-in volumeratio also dependent on the ratio between the specific heats at constantpressure and at constant volume, respectively, the K-value, of the usedelastic working fluid. This K-value varies between different gases whichmeans that the builtin pressure ratio is different in one and the samemachine in dependence on the gas used as working fluid.

One problem in a screw rotor machine of the type mentioned above is thatthe pressure within a chevronshaped chamber is not uniformlydistributed. In a compressor the pressure is thus highest at the tip ofthe chamber where pressure initiation takes place due to the penetrationof the land of the cooperating rotor into the groove. The pressure thusinitiated is then transmitted within the groove in the direction towardsthe end of the rotor. In the same way though not so accentuated thepressure is lower near the leading groove flank than near the trailinggroove flank as the gas enclosed in the groove is transported incircumferential direction as the rotors revolve. For this reason the gasenclosed in the groove is in a continuous movement in relation to theflanks of the groove.

When the chamber is opened towards the high pressure port the pressuredistribution in the groove is thus such that the pressure in certainportions thereof is higher than the pressure in the port while in otherportions it is lower than that in the port. The flow between the grooveand the high pressure port will then at the por tions of the groove inwhich there is an excess pressure in relation to the high pressure portbe in the direction from the groove to the port and at the portions ofthe groove in which there is a pressure lower than that in the highpressure port in the direction from the port to the groove. These flowscause an ununiform pressure distribution in the port resulting inturbulence and noise. Of course the flow conditions change due to thepressure equalization in the groove taking place during the continuedrevolving of the rotors. For each opening of a chamber towards the portthere is thus a periodic course resulting in an oscillation having afrequency coinciding with the number of the chevron-shaped chambersformed and opening towards the port per time unit, i.e. coinciding withthe product of the number of revolutions of the rotor and the number ofgrooves in the rotor.

The flows in different directions arising when the chevron shapedchambers open towards the high pressure port cause an abrupt decrease ofthe driving torque demanded by this chamber as the outflow from thepressure initiating zone is facilitated due to the sudden increase ofthe outflow area therefrom so that the initiating pressure decreases andas the pressure especially on the leading groove flanks increases. Thechanges in the driving torque caused in this way are so large that undercertain circumstances they cause a change of direction of the. torqueacting on the female rotor, the driving torque of which normally is onlya fraction of the corresponding torque of the male rotor, which meansthat the continuous driving contact between the elements transmittingtorque to the female rotor is broken and results in strong clearlynoticeable vibrations of the female rotor.

If these torque transmitting elements are the flank surfaces of the maleand female rotor lands the wear of the cooperating flanks isconsiderably increased.

On the other hand, if these torque transmitting elements are asynchronizing gearing connecting the male and female rotors thesevibrations can produce within the female rotor and especially within itsshaft portion carrying the, synchronizing gear secondary vibrationssubstantially of torsional type of such an amplitude that seizingbetween the rotors and corresponding wear-takes place. In such a waycomplete breakdowns of the rotors have occurred with synchronized aswell as with unsynchronized rotors.

The present invention has for its object to solve this oscillationproblem which, as stated above, substantially emanates from the pressurevariations in the high pressure port and the invention is based on thefact that it has been established by tests that no oscillations of thementioned type occur at a certain distribution of the driving torque ofthe machine between two cooperating male and female rotors.

By analysis of the frequencies of the noise existing in the outletchannel of a compressor and caused by pressure oscillations, which noisehas a sound level of about decibels, it has proved that most of thenoise has a frequency corresponding to the number per time unit ofemptied chevron shaped chamber, i.e. the product of the number ofrevolution of a rotor per time unit and the number of lands of therotor, or frequencies corresponding to harmonics thereof, which provesthat the pressure variations produced in a compressor outlet are of aconsiderable magnitude.

In the standard profile (embodying the symmetrical circular profiledisclosed in U.S. Patent No. 2,622,787) hitherto used the rotors havethe same outer diameters and the male rotor is provided with four landsand the female rotor with six lands. Each female rotor groove consists,when viewed in a transverse plane, of an inner portion limited by thepitch circle of the rotor and a circular are having its centre on thepitch circle and a radius which is 18% of the outer diameter of therotor, and of an outer portion limited by the circumscribing circle andthe pitch circle of the rotor and by addenda to the rotor lands limitingthe groove said addenda being located outside the pitch circle andhaving a radial extent corresponding to 14.3% of the length of the chordto the pitch circle lying within the land separating two adjacentgrooves.

With this standard profile theoretically about 6.7% of the driving poweris transmitted to the female rotor which owing to the speed ratiobetween the rotors means that the magnitude of the torque of the femalerotor is about 11% of the magnitude of the corresponding torque of themale rotor. By increasing the radial addendum of the female rotor itstorque may be increased so that at for instance an addendum having aradial extent corresponding to 20% of said chord about 9.5% of thedriving power is transmitted to the female rotor, i.e. the magnitude ofthe torque of the female rotor is increased to about 17% of themagnitude of the corresponding torque of the male rotor. At the sametime the displacement volume of the machine increases with about 3.5%which is also favourable.

However, an increase of the radial addendum of the female rotor alsoresults in effects of undesirable nature so that the magnitude of theaddendum must be limited and only for special applications can beincreased to such a magnitude as is shown for instance in the aforesaidPatent No. 2,622,787 where the addendum is completely circular whichmeans an addendum having a radial extent of about 50% of said chordinvolving that almost 24% of the power is transmitted to the femalerotor the torque of which will be about 46% of the corresponding torqueof the male rotor. The reason for this is among others that the blowhole between two adjacent chevron shaped chamhers caused 'by theaddendum will be too large in relation to the volume of said chambers.Thus, the flow through the blow holes results in that the pressure inall chevron shaped chambers is about equal to the pressure in the highpressure port of the machine which means a considerable decrease of theefficiency of the machine. Further, the increase of the powertransmission to the female rotor means an increased risk of wear of therotors especially in machineswithout synchronizing gears. Tests haveshown that a completely circular addendum brings about such a decreaseof the efiiciency that such addenda cannot be used in practice but inspecial types of machines where the characteristics obtained bycombination effects balance the decrease of the efficiency. A type ofmachine in which the circular addendum for this reason is used is a twostage tandem compressor where it is essential to bring about so largedifferences between the displacement volumes of the first and the secondstage that a pressure distribution between the stages can be obtainedwhich is most favourable as regards efficiency. Such a machine is shownin U.S. Patent 2,975,963.

In machines having a synchronizing gearing an increase of the powertransmission to the female rotor involves that the torque acting on thejournal shafts carrying the synchronizing gears and thus the risk oftorsional deformation of said shafts increases which in turn increasesthe risk of seizing between the rotors.

It has, therefore, been found suitable to shape the rotors in such a waythat the radial extent of the addenda of the female rotors lies withinthe range 18 to 37.5% of said chord which results in an increase of thetorque of the female rotor in relation to said standard profile by 25 to160% and a maximum increase of the blow hole by about 50% in relation tothe standard profile. This may be compared with the 50% addendumsuggested in the aforesaid U.S. Patent No. 2,622,787 which causes anincrease of the blow hole by about in relation to that of the standardprofile. It has proved that an optimum of the radial extent of theaddendum lies within the range 20 to 30% of the chord which gives anincrease of the torque of the female rotor by 40 to and a maximumincrease of the blow hole by about 33% in relation to the standardprofile. With respect to the shape of the corresponding rotor and thehousing it has been found especially suitable to give the addenum aradial extent of about 25% to the length of the chord which means anincrease of the torque of the female rotor by about 75% while theincrease of the blow hole is kept below 25% in relation to that of thestandard profile.

It has also proved that the risk of oscillations decreases with increaseof the built-in pressure ratio, i.e. with one and the same working fluidat a decrease of the area of the land flanks exposed to the pressurevariations in the high pressure port, in other words, the magnitude ofthe variations of the torque of the female rotor deriving therefrom inrelation to the driving torque transmitted between the rotors.

In practice it has thus proved that screw rotor machines with-rotors ofstandard profile operate perfectly without vibrations as compressorswith a built-in pressure ratio of about 7:1 which is used within a veryimportant field, viz. for producing compressed air for pneumatic toolsand the like which normally work at a pressure of about 7 kg./cm.

The main fields for screw rotor machines, however, require considerablylower built-in pressure ratios.

For compression of gases where it is of importance that the gas must notbe polluted by liquid, screw compressors are used which operate drywithout injection of cooling liquid into the working space. In suchmachines the practical upper limit of the built-in pressure ratio isabout 5:1 owing to the thermal deformation of the rotors and the housingin spite of the most intense cooling possible of the rotors and thehousing by a cooling fluid completely separated from the working fluidand passing through cooling channels.

Another important field for screw compressors is that for compression ofa cooling medium, especially for air conditioning plants. Compressorsfor this purpose are normally owing to the desired condensing andevaporating temperature designed for a built-in pressure ratio of about3:1 and usually for a built-in pressure ratio not exceeding 5 :1.

In practical use it has appeared that vibrations of the mentioned typearise in screw rotor machines designed for the built-in pressure ratiosactual in the fields mentioned above and provided with rotors ofstandard profile. This is true for machines operating dry as well as formachines with oil injection into the working space. It has thus beennecessary in order on the whole to bring machines of these types tofunction mechanically to provide them with synchronizing gearingsconnecting the rotors. Owing to the risk of seizing between the rotorsand damages of the rotors arising therefrom the clearances between therotors must be dimensioned under consideration of the angular deviationsof the female rotor from the synchronized position due to the arisingvibrations. This means that it has been necessary to make the clearanceslarger than if no vibrations in the rotors had occurred.

Tests have shown, however, that it is possible to prevent the generationof such vibrations in the female rotor also with built-in pressureratios lying below 5:1 by modifying the rotor profiles and thus thedistribution of the torque between the rotors in such a way that thetorque of the female rotor is proportionally increased. By elimi nationof such vibrations in the female rotor there is thus a possibility toreduce the clearances between the rotors while maintaining mechanicalreliability in operation and to obtain a reduction of the leakage losseswithin the machine which manifests itself by an increased efiiciency ofthe machine. This means that the screw rotor machine obtains anincreased capacity of competition with other known types of machinesused for the same purposes. At least in machines with oil injection intothe working space the advantage is furthermore obtained that the torquecan be transmitted between the rotors by direct flank contact which thusmeans an elimination of the synchronizing gearing and the costs thereforin connection with manufacture and adjustment at service as well as atthe manufacture of new machines which further influences the capacity ofcompetition of the machine in a favourable way.

In the following part of the specification the invention will bedescribed in detail in connection with a number of suitable embodimentsof the screw rotor machine shown in the drawings.

FIG. 1 shows a vertical section through a screw rotor compressor alongthe line 11 in FIG. 2.

FIG. 2 shows a cross section of the compressor shown in FIG. 1 along theline 22 in FIG. 1.

FIG. 3 shows another cross section of the compressor shown in FIG. 1along the line 3-3 in FIG. 1.

FIG. 4 shows a horizontal section of the compressor shown in FIG. 1 withthe rotors removed, the section being taken along the line 44 in FIG. 1.

FIG. 5 is a diagram showing the volume of a chamber when opening againstthe high pressure port plotted against the axial extent of the highpressure port.

FIG. 6 shows the cross section of a pair of cooperat' ing rotors mountedin the machine.

FIG. 6a shows on a larger scale the addendum of a female rotor landlocated outside the pitch circle of the rotor.

FIG. 6b shows on a larger scale the dedendum at the bottom of a malerotor groove located inside the pitch circle of the rotor.

FIG. 7 shows a vertical section through another screw compressoraccording to the invention.

FIG. 8 shows a vertical section through a third screw compressoraccording to the invention.

The screw compressor shown in FIGS. 1 to 4 comprises a casing 10, inwhich a working space 12 is provided in the shape of two intersectingcylindrical bores with parallel axes. The casing is further providedwith a low pressure channel 14 and a high pressure channel 16 for theworking fluid which communicate with the working space 12 through thelow pressure port 18 and the high pressure port 20, respectively.

In the shown compressor the low pressure port 18 is located entirely inthe low pressure end wall 22 of the working space 12 and substantiallyon one side of the plane containing the axes of the bores (FIG. 2). Inthe shown compressor the high pressure port is located partly in thehigh pressure end wall 24 and partly in the barrel wall 26 of theworking space 12 and entirely on the side of the plane containing theaxes of the bores opposite to the low pressure port (FIGS. 3 and 4).

In the working space 12 two cooperating rotors, one male rotor 28 andone female rotor 30, are located with their axes coaxial with the axesof the bores. The rotors are journalled in the casing 10 but for thesake of simplicity the bearings carrying the rotors are omitted. Therotors 28, 30 are further connected through a synchronizing gear 32 ofthe gear wheel type. The male rotor 28 is provided with an externalshaft 34 projecting from the casing 10.

The male rotor 28 is provided with four helical lands 36 withintervening grooves 38 which have a wrap angle of about 300. The femalerotor 30 is provided with six helical lands 40 with intervening grooves42 which have a wrap angle of about 200. As shown in FIG. 6 the flanksof the male rotor lands 36 are composed of three main portions, viz. anouter portion between the points 44 and 46 which at full intermeshbetween the male rotor land 36 and a female rotor groove lies inside thepitch circle 48 of the female rotor and which, when viewed in atransverse plane, is substantially circular around a centre located onthe pitch circle 50 of the male rotor with a radius which is 18.7% ofthe outer diameter of the male rotor, an intermediate portion betweenthe points 46 and 52 located between said outer portion and the pitchcircle 50 of the male motor, said intermediate portion being generatedby a point 54 on the flank of the female rotor groove 42 located at thepitch circle 48 of the female rotor, and an inner portion between thepoints 52 and 56 located inside the pitch circle 50 of the male rotorthe shape of which will be clear from the following portion of thespecification. The flanks of the female rotor lands 40 arecorrespondingly composed of two different portions, viz. an innerportion between the points 58 and 54 located inside the pitch circle 48of the female rotor and circular around a centre located on the pitchcircle 48 of the female rotor with a radius which is 18.7% of the outerdiameter of the female rotor, and an outer portion between the points 54and 60 located outside the pitch circle 48 of the female rot-or with aradial extent 62-60 the length of which is about 25.4% of the length ofthe chord to the pitch circle 48 of the female rotor lying Within afemale rotor land 40 and the end points of which are the point 54 and anequally located point 64 on an adjacent female rotor land flank. Theinner portion 52-56 of the male rotor land flank is shaped in such a waythat it continuously seals against the outer portion 54-60 of the femalerotor land flank, when said portion passes into or out of the dedendumof the male rotor groove 38 lying inside the pitch circle 50 of therotor. In order to make it possible to use smaller clearances betweenthe tips 44, 60 of the rotor lands and the barrel wall 26 of the casingand in this way to get a decreased leakage the tips of the rotor landsare shaped in such a way that the male rotor land 36 is slightly cutaway on both sides of its outermost portion 44 so that only a thinsealing edge is left, and the radially outermost portion 60 of theaddendum of the female rotor land 40 is also shaped as a thin sealingedge, a corresponding groove being provided at the innermost portion 56of the bottom of the male rotor groove 38.

Thus the female rotor has lands with profiles consisting of circularportions lying inside the pitch circle of the rotor with addendumportions connecting the circular portions and lying outside the pitchcircle of the rotor, and the male rotor has grooves with profiles havingcircular portions lying outside the pitch circle of the rotor connectedby dedendum portions at the bottoms of the grooves complementary inshape to the addendum portions of the female rotor lands.

The low pressure port 18 is so shaped that the rotor groovescommunicating therewith are open over their whole length when thecommunication is broken so that the complete displacement volume of themachine is always utilized. The high pressure port 20 which as saidabove comprises a portion located in the high pressure end wall 24 and aportion located in the barrel wall 26 is, however, shaped in such a waythat the rotor grooves 38, 42 when they start to communicate with thehigh pressure port 20 have a considerably smaller volume than thatcorresponding to completely open grooves. The high pressure port 20 isshaped in such a Way that its edges in the barrel wall 26 as well as inthe high pressure end wall are substantially parallel with the portionsof the lands limiting the grooves when the communication with thegrooves commences. In order to define the size of the high pressure portit is sufiicient to speak only about the position of the point 66 wherethe edge of the barrel wall sealing against the tip of the male rotorland 36 crosses of the point 66 from the low pressure end wall as isclearly shown in the diagram of FIG. Where the volume of the grooveportions forming a discharging chevron shaped chamber expressed inpercentage of the total volume of the two grooves is shown as a functionof the distance from the low pressure end wall of the point 66 expressedin percentage of the total distance between the low pressure and highpressure end walls.

The ratio of the total volume of the grooves to the volume of thechevron shaped chamber when opening against the high pressure port iscalled the built-in volume ratio of the machine, 6. The mean pressure inthe chevron shaped chamber when it opens against the high pressure portcan be calculated by means of the built-in volume ratio and theIc-VEtlLlB of the working fluid defined already in the introduction tothe specification. The ratio of the mean pressure in the chevron shapedchamber when it opens against the high pressure port to the pressure inthe grooves completely filled by working fluid when communicating withthe low pressure port, the built-in pressure ratio can thus becalculated from the formula The machine operates in the following way.The rotors are rotated by a motor (not shown) through the shaft 34projecting from the casing 34 and through the synchronizing gear 32connecting the rotors. Working fluid is sucked in through the lowpressure channel 14 and the low pressure port 18 into the Working space12 where it flows into the portions of the rotor grooves 38, 42 whichare open against the low pressure port 18 so that the grooves are filledwith working fluid to their whole length. The working fluid is thenmoved circumferentially 'by the rotors. During this movement a femalerotor land 40 first enters the male rotor groove 38 and a male rotorland 36 then also enters the female rotor groove 42. In this way achevron shaped chamber composed of two communicating grooves is formedwhich is sealed against the high pressure port 20 as well as against thelow pressure port 18 and the volume of which continuously decreasesduring the rotation of the rotors. During this decrease of the volumethe pressure in the chamber increases in the first hand at the intermeshbetween a land and the cooperating groove. The pressure initiated inthis way is then transmitted through the rotor grooves in the directiontowards the high pressure ends of the rotors. Owing to the rotation ofthe rotors a composed axial and circumferential movement is imparted tothe mass of working fluid enclosed in each groove where the axialcomponent dominates near to the apex of the chevron shaped chamber andthe circumferential component dominates near to the base of the chevronshaped chamber. In this way there is a higher pressure in the workingfluid adjacent to the driving trailing flank of the rotor then adjacentto the leading flank of the rotor. The pressure distribution within achevron shaped chamber is thus very ununiform but becomes more and moreuniform as the rotors revolve and the pressure wave from the intermeshis transmitted from the apex towards the base of the chevron shapedchamber. In working fluids in which the velocity of sound is relativelyhigh this pressure transmission also occurs faster than in workingfluids in which the velocity of sound is relatively low. After a certainrotation of the rotors corresponding to a predetermined turning angle ofthe male rotor the chevron shaped chamber is opened towards the highpressure port. The pressure distribution within the chevron shapedchamber is then still ununifonm so that the pressure in the portionthereof located most adjacent to the apex of the chamber is higher thanthe pressure in the high pressure channel while the pressure in otherportions of the chamber more adjacent to its base is lower than that inthe high pressure channel. Owing to this fact a flow condition composedof ditferent partial flows in PP S 'E directions arises when the chamberopens against the high pressure port. Owing to the high pressure closeto the apex of the chamber and the sudden increase in the out flow areathrefrom a substantial flow arises in this portion of the high pressureport in the direction from the chamber to the high pressure channel sothat the pressure in said portion of the chamber decreases which resultsin an immediate reduction of the power with which the rotors act on theworking fluid, i.e. a rapid decrease of the torque necessary forrotating the rotors. At the same time a flow arises in the portion ofthe port which is closest to the base of the chevron shaped chamber inthe direction from the high pressure channel to the grooves so that thepres-sure in the grooves increases rapidly to the same level as thepressure in the channel. In this way the pressure in the groovesincreases especially adjacent to the leading flanks which causes a pushtrying to drive the rotors forwardly with a higher speed than that atwhich they normally rotate. Also in this way a rapid decrease of thetorque necessary for rotating the rotors is obtained. The total decreaseof the torque thus obtained may lead to that the female rotor for amoment is driven by the ga instead of driving the gas if the femalerotor has such a profile that its power transmission is below a certainpercentage of the total power transmission. When the rotors continuetheir rotation the pressure in the grooves is rapidly equalized exceptfor the zone closest to the intermesh from where, however, the out flowarea is so large that the pressure at this point does not at all reachthe same level as it the intermesh point had still been covered bybarrel walls surrounding the rotors. The working fluid flows in this wayfrom the working space 12 through the high pressure port 20 to the highpressure channel 16- and from there out from the machine.

Owing to the above flow conditions there is a periodic variation of thetorque load of the rotors which variation has a frequency whichcorresponds to the number of chevron shaped chambers opening to the highpressure port per time unit or, in other words, the product of thenumber of revolutions of the rotor and the number of the lands of therotor. By shaping the profiles of the rotors in accordance with theinvention it is guaranteed that the torque driving the female rotorforward is greater than the decrease thereof arising when a chamberopens to the high pressure port, whereby it is avoided that the rotor issubjected to an acceleration at this moment causing vibrations of therotor.

The compressor shown in FIG. 7 differs from the one shown in FIG. 1 onlyby the fact that the synchronizing gear 32 connecting the rotors hasbeen omitted so that the torque transmission between the rotors isobtained by direct flank contact therebetween.

The compressor shown in FIG. 8 differs from the one shown in FIG. 1 onlyby the fact that the synchronizing gear 32 connecting the rotors hasbeen omitted so that the torque transmission between the rotors isobtained by direct flank contact therebetween, and by the fact that thecompressor is provided with openings 68 in the casing 10 for injectionunder pressure of a liquid supplied from a pressure liquid source, notshown, close to the intersection line between the bores of the workingspace 12 on the same side of the plane of the axes of the bores as thehi h pressure port 20, whereby on one hand an improved cooling andsealing is obtained within the compressor and on the other a liquid filmis obtained on the rotor flanks contacting each other.

The invention is of course not limited to the embodiments shown butcomprises everything falling within the scope of the following claims.

What I claim is:

1. A screw rotor machine for elastic working fluid comprising twocoplanar intermeshing male and female rotors having helical lands with awrap angle of less than 360 and intervening grooves and a casingenclosing said rotors in sealing relationship to provide working spacefor the fluid, said lands having profiles comprising substantiallycircular portions on the male flanks lying outside the pitch circle ofthe male rotor and substantially circular portions on the female flankslying inside the pitch circle of the female rotor, each of the lands ofthe female rotor further having an addendum portion extending outsidethe pitch circle of the female rotor, the radial extent of said addendumbeing within the range of from 18% to 37.5% of the length of the chordof the portion of the pitch circle of the female rotor lying Within theland and the bottom portion of each of the grooves of the male rotorhaving a dedendum portion lying inside the pitch circle of the malerotor and having a profile complementary to that of said addendumportions of the female rotor land.

2. A screw rotor machine as defined in claim 1, in which the radialextent of the addendum portions of the female rotor lands is within therange of from 20% to 30% of the length of said chord.

3. A screw rotor machine as defined in claim 2, in which the radialextent of said dedendum portion is approximately 25% of the length ofsaid chord.

4. A screw rotor machine as defined in claim 3, in which the male rotorhas four lands and the female rotor has six lands with interveninggrooves and in which the portion of each female rotor groove locatedinside the pitch circle of the rotor has a profile comprised by acircular arc having its center on the pitch circle of the female rotorand a radius the length of which is approximately 18.7% of the outerdiameter of the rotor, whereby to provide male and female rotors ofapproximately the same outer diameter.

5. A screw rotor machine as defined in claim 1 in which the rotors andthe high and low pressure ports communicating With the working space areso formed and dimensioned in relation to each other that the built-inpressure ratio for the elastic working fluid is less than 5 to 1.

6. For use in a screw rotor machine of the character described, a pairof cooperating rotors, comprising a male rotor and a female rotor havinghelical lands with a wrap angle of less than 360 and interveninggrooves, said lands having profiles comprising substantially circularflank portions on the male flanks lying outside the pitch circle of themale rotor and substantially circular portions on the female flankslying inside the pitch circle of the female rotor, each of the lands ofthe female rotor further having an addendum portion extending outsidethe pitch circle of the female rotor, the radial extent of said addendumportion being within the range of from 18% to 37.5% of the length of thechord of the portion of the pitch circle of the female rotor lyingwithin the land and the bottom portion of each of the grooves of themale rotor having a dedendum portion lying inside the pitch circle ofthe male rotor and having a profile complementary to that of saidaddendum portions of the female rotor lands.

7. Rotors as defined in claim 6 in which the radial extent of saidaddendum portions is within the range of from 20% to of the length ofsaid chord.

8. Rotors as defined in claim 7 in which the radial extent of saidaddendum portions is approximately 25% of the length of said chord.

9. Rotors as defined in claim 8 in which the male rotor has four landsand the female rotor has six lands with intervening grooves and theportions of the female rotor grooves lying inside the pitch circle ofthe female rotor are defined by profiles comprised of circular arcshaving their centers on the pitch circle of the female rotor and aradius the length of which is approximately 18.7% of the outer diameterof the female rotor.

References Cited by the Examiner UNITED STATES PATENTS 2,486,770 11/1949Whitfield 103128 2,622,787 12/1952 Nilsson 230-143 2,975,963 3/1961Nilsson 230-l43 3,179,330 4/1965 MacColl 230--143 MARK NEWMAN, PrimaryExaminer.

WILBUR J. GOODLIN, Examiner.

1. A SCREW ROTOR MACHINE FOR ELASTIC WORKING FLUID COMPRISING TWOCOPLANAR INTERMESHING MALE AND FEMALE ROTORS HAVING HELICAL LANDS WITH AWRAP ANGLE OF LESS THAN 360* AND INTERVENING GROOVES AND A CASINGENCLOSING SAID ROTORS IN SEALING RELATIONSHIP TO PROVIDE WORKING SPACEFOR THE FLUID, SAID LANDS HAVING PROFILES COMPRISING SUBSTANTIALLYCIRCULAR PORTIONS, ON THE MALE FLANKS LYING OUTSIDE THE PITCH CIRCLE OFTHE MALE ROTOR AND SUBSTANTIALLY CIRCULAR PORTIONS, ON THE FEMALE FLANKSLYING INSIDE THE PITCH CIRCLE OF THE FEMALE ROTOR, EACH OF THE LANDS OFTHE FEMALE ROTOR FURTHER HAVING AN ADDENDUM PORTION EXTENDING OUTSIDETHE PITCH CIRCLE OF THE FEMALE ROTOR, THE RADIAL EXTENT OF THE ADDENDUMBEING WITHIN THE RANGE OF FROM 18% TO 37.5% OF THE LENGTH OF THE CHORDOF THE PORTION OF THE PITCH CIRCLE OF THE FEMALE ROTOR LYING WITH IN THELAND AND THE BOTTOM PORTION OF EACH OF THE GROOVES OF THE MALE ROTORHAVING A DEDENDUM PORTION LYING INSIDE THE PITCH CIRCLE OF THE MALEROTOR AND HAVING A PROFILE COMPLEMENTARY TO THAT OF SAID ADDENDUMPORTIONS OF THE FEMALE ROTOR LAND.